Transmission gearing



July 29, 1941. F.` w. coTTERMAN 2,250,889l

TRANSMISSION GEARING Filed Jan. 11, 1940 4 Sheets-Sheet 1 uw mw.. ww

. WEA/rok' July 29, 1941- F. w. cor'rERMA 2,250,889

TRANSMISSION GEARING Filed Jan. 11, 1940 4 Sheets-Sheet 2 Filed Jan. 11,1940 4 Sheets-Sheet 5 July 29 1941. F. w. coT'rERMAN TRANSMISSIONGEARING 4 sheets-sheet 4 Filed Jan. 11, 1940 Hillllllll adp lll/[MagPatented July 29, 1941 UNITED STATES PATENT OFFICE TRANSMISSION GEARINGFrederick W. Cotter-man, Dayton, Ohio, assignor of one-half to Bessie D.Apple, Dayton, Ohio Application January 11, 1940, Serial No. 313,375

(rc1. P14-26o) 15 Claims.

`This invention relates Y'to transmission gearing and is particularlyapplicable to motor vehicles, part of the mechanism employed beingsubstantially shown in my copending application Serial No. 257,059,iiled February 18, 1939.

An object of the invention is to provide transmission gearing having sixforward gear ratios with the highest ratio an overdrive, together withclutch mechanism for connecting the engine to I the transmission, allfully automatic and occupying no greater space than commercial handshift devices of equal range and capacity.

Another object is to so construct and arrange the transmission gearingand clutch mechanism that if the mechanism is operating in eitherunderdrive, direct, or overdrive ratios and there arises a need for morepower than the engine can deliver at the then existing speed, a stepdown in ratio will automatically take place to allow the engine to riseto a more appropriate speed.

Another object is to provide a transmission gear set comprising a'sungear, a ring gear, planet pinions and carrier, with speed responsiveclutch means ,to connect the several elements variously between theinput and output members to provide underdrive, direct and overdriveratios, and a booster gear set also comprising a sun gear, a ring gear,planet pinions and carrier, with means responsive to both speed andtorque to connect the booster gear 'setin series with the transmissiongear set, whereby a step down or a step up of one speed is always hadwhen speed-load conditions warrant, no matter in which of its severalratios the transmission gear is then operating, whereby the device isalso subject to the will of the driver, in that he may by suddenlychanging the amount of applied power by means of the engine accelerator,cause a shift up or down as the case may be.

Another object is to provide the transmission gear set with two positiveclutches, the rst being on the transmission output member and normallyclutching the carrier and the second on the transmission input memberand normally clutching the ring gear, whereby the transmission gear setacts as a speed reducing device or underdrive, the rst clutch being`operable at a relatively low predetermined speed to release the carrierand clutch the ring gear, whereby both input member and output memberwill be clutching the ring gear to provide a direct drive ratio, and thesecond clutch being operable at a higher predetermined speed to releasethe ring gear and clutch the carrier, whereby the gearing actsas a speedincreasing device or overdrive, the sun gear being at al1 times xedlysecured against rotation.

Another object is to so construct the clutch mechanism of thetransmission gear set that there will be positive two-direction drivingAconnections between the several elements, in underdrive, in directdrive, and in overdrive, and so that when a shift from one ratio to theother is taking place, the clutches, both of which are operative torelease one element and clutch a second, always clutch the said secondbefore they release the rst, to the end that there will be no freewheeling, either in underdrive, direct, or overdrive, or during thetransition period in the shift from any one ratio to another.

Another object is to provide a main engine clutch, responsive to thespeed of the engine, to connect the engine to the transmission inputmember through the booster gear, and an auxiliary engine clutch,responsive to the speed of the vehicle to connect the engine to thetransmission input member directly and independently of the boostergear, whereby, if the vehicle is coasting while the engine is dead or isidling, the engine will be connected for engine braking at a lowvehiclespeed by said auxiliary clutch. Another object is to so constructand arrange the gear mechanism with respect to the main and auxiliaryvengine clutches, that the engine clutches will be contained nanentirely separate housing from the gears, whereby the gears maybekfullyY lubricated and the engine clutches may be kept dry, to the endthat dry plate clutches, which have been proven the most adaptable, maybe employed.

Another object is to so construct the main engine clutch that itsengagement secures the ring gear of the booster gear set to the engine,then connect the carrier to the transmission input member' and provide aone way brake to prevent backward rotation of the sunV gear, to the endthat, whenever the main clutch engages, 'engine power at' reduced speedwill be transmitted thru the ring gear to the carrier and therefore tothe transmission input member.

Another object is to provide, inI the booster gear set, gearing withhelical teeth, so angled that the tangential load carried by the gearingcauses an end thrust in a direction proper for disengaging the auxiliaryengine clutch, with proper means to apply the end thrust to theauxiliary engine clutch to disengage't and keep it fully disengaged aslong as the booster'gear Vmitted by said gear.

is transmitting power, to the end that no dragging action may be presentin the auxiliary clutch by partial or insufficient engaging pressure.

Another object is to provide, in both the engine clutches and thetransmission clutches, means for causing the centrifugal weights of aset to all move together, to the end that no one weight of a set maymove outwardly ahead of the others and thereby cause an unbalancedeffect.

Another object is to so arrange the connection between the main engineclutch and the booster ring gear that, altho the ring gear is connectedto be rotated, it may nevertheless move axially by load on its helicalteeth, to the end that the axial pressure of the ring gear which willvary with the torque being transmitted, may delay engagement of theauxiliary engine clutch which eliminates the drive thru the boostergear.

Another object is to so construct the auxiliary engine clutch that it isnormally disengaged, and place its speed responsive mechanism on avehicle driven member, whereby, Starting of the vehicle from rest `willalways be done thru the booster gear, altho the length of time thebooster gear will continue in eifect will depend on the balance betweenthe end thrust of the ring gear and the vehicle driven speed responsivemeans of the auxiliary clutch.

Another object is to provide for the auxiliary engine clutch, whichengages to eliminate the booster gear, a resilient means normallyinoperative to engage the clutch, and centrifugal weight means rotatedin proportion to vehicle speed and normally adapted, at a loW vehiclespeed, to first apply said resilient means to urge engagement of saidclutch, then further stress the resilient means to more strongly urgeclutch engagement as the vehicle speed increases, whereby the speed atwhich the axial thrust of the booster ring gear may be overcome, and thebooster gear eliminated, will vary with the torquebeing trans- Anotherobject is to so construct the resilient means and the centrifugal weightmeans of the auxiliary engine clutch that the force of the weights willbe applied to stress the resilient meansthru a leverage which becomesprogressively less effective as the speed increases, Whereby the stressof the resilient clutch engaging means will increase at a rate which isless than directly proportional to the R. P. M. instead of at a rateproportional to the square of the R. P. M. las'it does where the forceof centrifugal weight means is applied directly, or thru an unvaryingleverage, as in common practice, to the end .that sufficient clutchengaging pressure may be had at the lower speeds without having toogreat a `clutch engaging pressure at the higher speeds.

Another object is to so construct the clutch mechanism which controlsthe booster gear that direct drive will always be fully accomplishedbeforebooster gear drive is eliminated, the one, by engagement, liftingthe load off the other, to the end that there Awill be no period betweenbooster gear drive and direct drive in` which there is no drive, asthere is in conventional gear shift all other times and under all otherdriving conditions.

Another object is to so construct and arrange the ratio changingmechanism, that, altho acceleration may be effected by shifting from lowthru second to high, it may, whenever desirable, be effected by shiftingdirectly from low to high.

Another object is to so proportion and arrange the speed responsivedevices which shift the transmission clutches that, altho they aredesigned to shift from one ratio to the next at certain fixed speeds,they will not do so as long as power is being transmitted thru the ratiowhich is then in effect, then provide a signalling device which becomesoperative to show when a shift should preferably be made and againinoperative to show when it has been made.

Another object is to provide a manually operable lockout means wherebythe fourth or overdrive ratio may be rendered Wholly inelfective,so-that, when driving in mountainous country, the highest ratio will bethird speed or high, both for forward driving and for engine braking.

That the foregoing enumerated objects and other meritorious features areattained in the embodiment of the invention hereinafter illustrated anddescribed will be apparent when the specification is read with referenceto the drawings, wherein,

Fig. 1 is a longitudinal axial section taken thru the transmission as atI--l of Fig. 2.

Fig. 2 is a transverse section, taken at 2-2 of Fig. 1, showing theconstruction of the main and auxiliary engine clutches and the boostergear set principally in elevation.

Fig. 3 is a half transverse section, taken at 3--3 of Fig. 1, showingthe roller brake which is provided for holding the ring gear of thebooster gear set from rotating backwardly.

Fig. 4 is a half transverse section, taken at -d of Fig. 1, thru thehinge and work arm of one of the centrifugal weights which effect ashift in the transmission gear set from underdrive ratio to direct driveratio.

' Fig. 5 is a half transverse section, taken at 5-5 of Fig. 1, thru thebody of one of the underdrive-to-direct centrifugal weights, and thruthe jaw clutch which normally connects the planet pinion carrier of thetransmission gear set to thel transmission output member.

Fig. 6 is a half transverse section, taken at 6-6 of Fig. 1, showing thegearing of the transmission gear set.

Fig. 'Z is a transverse section, taken at 1-1 of Fig. 1, thru thesignalling device which is provided to show when any shift intransmission ratio is desirable and when it has been made, and thru thejaw clutch which is normally disconnected but is operable at a certainspeed to connect the transmission input shaft to the planet pinioncarrier of the transmission gear set to effect overdrive ratio,

Fig. 8 is a plan view of the main parts of the direct-to-overdrive jawclutch, one of the parts of which appears in cross section in Fig. 7.

Fig. 9 is a fragmentary transverse section, taken at 9 9 of Fig. 1,thruthe manually operable mechanism which sets the transmission forforwardneutral or reverse.

, Fig. 10 is a part transverse section taken at lll-,l0 of Fig. 1, thruthe hinge and work arm of one of the centrifugal weights which effect ashift in the transmission gear set from direct drive ratio to overdriveratio.

Fig 11 is a transverse section, taken at H--II of Fig. 1, thru theshifting fork of the lockout mechanism which eliminates the overdrive,and thru the detent mechanism which holds the lockout and the forward,neutral and reverse mechanism in the selected position.

Fig. 12 is a transverse section taken at |2|2 of Fig. 1, thru theforward, neutral and reverse shifting fork.

Fig. 13 is a half transverse section thru the reversing gear set andthru the shifting mechanism which operates the overdrive lockout.

Fig. 14 is a transverse section thru the speedometer gearing.

Fig. 15 is a diagram showing the action of the centrifugal weights ofthe auxiliary engine clutch as the weights swing outwardly about theirhinge pins to dierent angular positions, the diagram giving the totaldeflections of the clutch engaging springs at each unit of angularmovement of the weights, the stress of the springs in each weightposition and the R. P. M. required of the weights to produce inthesprings the given clutch engaging stresses thru the leverages available.

Fig. 15 is a curve chart, plotted from the values in Fig. 15, showingthe available engine power at any speed, and what portion of the poweravailable at any given speed may be applied at that speed withoutbringing the speed-torque controlled booster gear into action, thebooster gear coming into action with progressively less powerapplication as the engine speed is lower and the torque output less. Italso shows, by comparison, the difference in result obtainable whencentrifugal force is applied thru a progressively less effectiveleverage as compared with direct application.

Fig. 1'7 is a plan view ofthe main parts of the underdrive-to-direct jawclutch, one of the parts of'which appears in cross section in Fig. 5..

Fig. 18 is a wiring diagram showing the connections for the electricalsignal which indicates when shifting of the clutches should occur.

Construction 'I'he clutch housing 26 may bel secured to the engine 28 byscrews 21. A booster gear housing 29 is formed integral with the clutchhousing by depressing the rear wall thereof. The transmission gearvhousing 30 is secured to the clutch housing by the screws 3|. Apartition 32 is interposed between the open ends of the booster gearhousing 29 and the transmission gear housing 30. The reverse gearhousing 33 is integral with the transmission housing 30, a partitionwall 34 separating them. The rear bearing head 35 is held to the housing33 by screws 36.

Secured to the crankshaft 31 by bolts 38 is the flywheel 39, the rim 40of which has internal splines 42 to which the external splines of themain clutch backing plate 43 and pressure plate 44 are slidably tted. Aspring ring 45 in a groove in the rim 40 limits forward movement of thebacking plate 43.

The main clutch frame 46 is secured to the flywheel rim 40 by screws 48and carries a series of hinge ears 49 (see Fig. 2) to which the mainclutch weights 50 are swingably held by hinge pins 52. Pressure plate 44has a series of pins 53 which extend thru holes in the frame 46, theends of the pins touching the upper front face of the weights.

A second series of pins 54 carried by the pressure plate 44 have theirrear ends bearing against the lower front face of the weights. Midwaybetween adjacent weights 50 are a series of hubs 55. Pressure plate 44has a series of studs 56 extending thru the frame 46 and iitting itclosely but slidably. Y

'I'he hubs 55 are counterbored to receive the springs 58. Collars y59held onthe free end of the studs 56 by nuts 60 fit the counterbored partof the hub closely but slidably andv hold the springs 58 under aninitial tension. The close tting studs 56 and collars 59 serve as guidesto restrain one side of the pressure plate 44 moving ahead of the otherand consequently cause the weights 50 to move out in unison.

The clutch plate 62 is faced with a commercial dry clutch facing 93. Theinner diameter of the plate isvflanged at 64 and carries the studs 65and rollers 66 thru which the plate transmits its power when clampedbetween the backing plate 43 and pressure-plate 44. Themain clutch maybe broadly designated by the numeral 10.

The transmission input shaft 63 has external splines 69 over which theinternally splined hub 12 of the auxiliary clutch frame 13 is tted.

The clutchframe 13 is provided with pairs of hinge ears 14 between whichthe weights 15 of the auxiliary clutch (see Fig. 2) are swingablysupported by the hinge pins 16. Each weight has a pair of hubs 18 on areduced outer endof each of which a roller 19 is rotatable. The rollersare held in place by washers 89 which are held on the reduced end byriveting. The ears 14 are so shaped on their outer edges as to provide astop for the hubs -19 to limit inward swinging of the weights. f l 1 Theauxiliary-clutch pressure plate 82 has a hub 83 slidableaxially over thehub 12 ofthe clutch frame. A light spring 5| always urges plates 13 and62 axially-apart: A series of guide studs 84 are heldv angularly spacedin the pressure plate 82 by the nuts-85. The studs 94Y may be hollowedfor lightness. Y

Y A spring compressing plate 86 has a series of hubs 88 extendingforwardly intermediate the ears 14, and a series of arms 89 extendingoutwardly from the hubs, eacharm 09 lying immediately in back of and incontact with a roller 19. The hubs 88 are bored at their outer ends tofit over `the guide studs 84 closely but slidably, then counterbored toreceive the springsy 92, the enlarged outer ends 93 of the studs 84being slidably tted to the counterbores. The clutch plate 94 is facedwith linings 95 similar to the main clutch plate 62 and has externalteeth 96 which lit slidably into theinternal splines 98 of the flywheelrim 49. The auxiliary clutch may be broadly designated by the numeral|00.

The booster gear set which is contained in the housing 29 and enclosedtherein by the partition 32, comprises a planet pinion carrier 99, thehub |02 of which is internally splined to fit over the external splines91 of the shaft 68. The carrier Y 99has a series of angularly spacedstuds |04 each of which has rotatable thereon a planet p-inion |05provided with a bearing bushing |06. A washer I0| and rivet |03 holdseach pinion from axial movement.

i A sun gearv |08 and .the inner member I I0 of a roller brake are endsplinedtogether at I 2 I, both being press tted'over the same bearingbushing |09 which is rotatable on the outside oi the hub |02. The outerring II3 is concentrically held to the partition plate 32 by screws 90.Rollers I I2 cooperate with the inner member IIO and outer ring I I3 tohold the sun gear from rotating back'- wardly, the usual springs III andplungers II-'l being provided to urge the rollers toward operativeposition. By backwardly is meant anticlockwise when viewed from the leftof Fig. 1. The sun gear |08 is in constant mesh with the planet pinions|05. The roller brake may be broadly designated by the numeral |01 (seeFig. 3)

v The ring gear I4, also in mesh with the planet pinions |05, has aforwardly extending hub H5 which has press tted therein a bearingbushing H6 which is freely rotatable on the transmission input shaft 68.A ring gear driving member H8 has a rearwardly extending hub I9 which isslidably fitted over the bearing bushing I I6.

The hubs H5 and H9 are end splined together at and a spring ring |21,snapped into a groove in the outside ci the bushing, holds the hub H9 onthe bushing. The rear end of the bushing is enlarged to slidably fitinto a counterbored opening 9| in the hub |02. A small hole 81 connectsthe counterbore 9| to the hollow interior of the shaft 68, whereby theenlarged end of the bushing may have a dash pot action in moving axiallyin the counterbore 9|, its rate of movement being limited by the flow ofoil thru the small hole 81, thereby preventing a too rapid engagement ordisengagement of .the clutch |90.

An annular groove 8| collects oil escaping forwardly between shaft 68and bushing H6 and returns it to housing 29 thru small hole 11. A11 oilthrow rib 1| on shaft 68 assists in throwing 01T the leakage oil. Asecond annular oil collecting groove 61 further forward in bushing H6has holes 6| which connect to notches 51 to slightly lubricate thethrust bearing |32,

The booster gear housing 29 has a forwardly extending hub |22 providedwith a bearing bushing |23 in which the hubs H5 and I9 are runninglyfitted. The forward end of the hub |22 is enlarged to contain theannular groove |24 which catches any oil escaping from the end of Athebushing |23 and returns it thru the hole |25. An oil throw |28, formedon the ring gear driving member H8 assists in conning the leakage oil tothe groove |24. A second annular groove v|29 is formed in the ring geardriving member H8, this second groove being provided to collect any oilwhich may escape past grooves 8| and 61 and thrust bearing |32. Smallholes |38 are provided to transfer any oil collected in groove |29 tothe groove |24.

The end thrust bearing |32 is preferably made of Igraphite impregnatedbearing metal such as is now commercially available for clutch thrustbearings. The thrust bearing |32 is brought to a sharp outer edge at |33to assist it in throwing any drops of oil which may reach it, into thegroove |29.

A circular row of shouldered pins |34 are secured in the ring |32 andare freely slidable thru holes in the auxiliary clutch frame 13, theirends normally bearing against the end of the hub 83 of the auxiliaryclutch pressure plate.

The ring gear driving member H9 has a rim |35 the outside of which isprovided at suitably spaced intervals with slots |36 which extendentirely thru the rim. Slots |36 fit over the rollers 66 closely butrunningly, whereby the driving member I i8 may shift axially withrespect to the clutch plate 62 while under load.

A shoulder |38 on the splines 91 secures .the carrier 99 against axialmovement on the shaft 68. A small bronze washer |39 takes any slightrearward end thrust which the carrier may have, the carrier being, ofcourse,` balanced against axial movement between the axially rearwardthrust on the sun gear and the axially forward thrust on the ring gear.'Ihe sun gear needs no thrust washer inasmuch as it never rotates Whileunder load.

A bronze washer |40 limits forward movement of the ring gear to theposition shown, its rear ward movement being arrested when the space |4|is taken up. A thrust washer |42 as Well as the bearing bushing |43which rotatably supports the forward end of the shaft 68 may preferablybe made of graphite impregnated bearing metal.

The rear end of the shaft 68 is rotatable in a bearing bushing |44 pressfitted into the hub of the transmission output member |45.

Midway the partitions 32 and 34 inthe housing 30 is the transmissiongear set which provides an underdrive, a direct, and an overdrive ratio.The sun gear |46 has a long bearing bushing |48 press fitted therein,the transmission input shaft 68 being runningly fitted in this bushing.A hub |52 extends rearwardly from the partition plate 32 and a bushing|54 is press fitted into this hu-b. 'I'he sun gear |46 and the hub |52of the partition member are end splined together at |56 whereby the sungear is positively held against rotation at all times.

The Planet pinion carrier of the transmission gear set comprises a frontbearing member |58 provided with a bearing bushing |60, and a rearbearing member |62 provided with a bearing bushing |64. Planet pinionbearing hubs |86 hold the carrier bearing members axially spaced apart,and the bolts |68 and nuts |10 extending thru the carrier bearingmembers and the pinion lbearing hubs hold the carrier parts together.

Planet pinions |12 having bearing bushings |14 are rotatable on thebearing hubs |66, the pinions being in constant mesh with the sun gear|46.

The ring gear |16 is in constant mesh with the planet pinions 12. Itsfront bearing member |18 and its rear bearing member |80 are secured tothe ring gear by bolts |82 and nuts |84. The front bearing member |18 isprovided with a bearing bushing |86 and the rear bearing member with abearing Ibushing |88.

These bearing bushings enable the ring gear to rotate in concentricrelation with the sun gear, but carry no radial load except the weightof the several parts.

The outpult member |45 of the transmission gear set has arearwardlyextending hub |90 rotatable in the ball bearing |92, held in thepartition 34, the'front end being closed by the bearing head |94 securedin place by the screws |96. The bearing head |94v is provided with abearing bushing |98. End thrust washers |95, |91, |99, 28| and 203 limitaxial movement of the several parts.

For convenience in further description, the ring gear |16, its bearingheads |18 and |80, its bolts |82, and nuts |84 and its bearing bushings|86 and |98 may be collectively referred to as the ring gear element.For the same reason, the planet pinion carrier front bearing member |58and rear bearing member |62 with their bearing bushings |69 and |84, andthe planet pinion'bearing hubs |66 with their bolts |68 and nuts |10 maybe collectively referred to as the carrier element. c v

Obviously, with the sun gear |46 permanently held from rotating by theend splines |56 as hereinbefore described, if the ring gear element isrotated, the Icarrier element will rotate in the same direction but atless speed, and if the carrier element is rotated, fthe ring gearelement will rotate in the same direction but at greater speed. The ringgear element will under all conditions, rotate faster than the carrierelement.

It follows that, if the input member of the transmission gear set isconnected to the ring gear element, and the output member to the carrierelement, an underdrive ratio will be provided wherein the output memberwill rotate more slowly than the input member.

Conversely, if the input member is connected to the carrier element, andthe output member to the ring gear element, an overdrive ratio will beprovided wherein th-e output member will rotate faster than the inputmember.

On the other hand, if both the input member and the output member areconnected at the same time to the same element, a direct drive will beprovided wherein the input member and output member revolve at the samespeed. Both members in this case may preferably be connected to the ringgear element for then the carrier element merely rotates idly at subengine speed as ydoes the countershaft of a conventional synchromeshtransmission during direct drive.y

Of course, a direct drive may be had by connecting the input member andthe output member both at the same time to the carrier element, but inthat. case the ring gear element will rotate idly at super-engine speed,which is less desirable. 4 v

It will now be apparent that, with the single planetary gear train,arranged as shown, an underdrive ratio, a direct drive ratio, and anoverdrive ratio may beV had by providing the input and output memberseach with a clutch which will, each at its own proper time, take hold ofone of the rotating elements, i. e., ring gear element o1- carrier'element, and let go of the other,

Accordingly two clutches are provided. The clutch which is carried onthe output member has one jaw member which normally engages jaws on thecarrier element, and a second normally idle jaw member which may becomeoperative labove a predetermined speed to rst engage jaws on the ringgearV element, then cause the first jaw member to release, the jaws onthe carrier element.

`The clutch which is carried onthe input member has one jaw member`which normally engages jaws on the ring gear element, and a secondnormally idle jaw member which may become operative above a higherpredetermined Speed, to first engage jawson the carrier element, thencause the first jaw memberto release the jaws on fthe ring gear element.

The clutch which is carried bythe output member, and whichfunctions toshift from an underdrive ratio to a direct ldrive ratio may be forgreater convenience -in further description called the direct driveclutch. The other clutch which is carred by the input member, and whichfunctions to shift from a direct drive ratio to an overdrive ratio, mayfor a like reason be called the overdrive clutch.

The direct drive clutch is carried in the output member bearing head l94which has internal splines 288. Jaw members 2l0 and 2|2 have externalsplines which are axially slidable in rv On the outside of the head |94,eight weights 2 I4 are equally spaced and hinged between ears 2|8 byhinge pins 2|8, the work arms 220 of the weights extending thru slots222 in the head. The outward movement of the weights is accuratelylimited by contact of the work arms 228 with the# rear edges of theslots.

The inner ends of the work arms 228 extend into an annular groove in theshift collar 224 which also has external splines slidably fitted to theinternal splines 208. A Weight return spring washer 225 has externalsplines also slidable in the internal splines 208, and a weight returnspring 228 under considerable stress urges the washer 226 forwardly.

In the instant embodiment there are twelve equally spaced internalsplines 288 and on the collar 224 and washer 226 there are twelveexternal splines slidable in the internal splines. On the jaw members2H) and 2I2, however, every fourth external spline is cut away (see Fig5) which leaves three spaces between internal splines 288 into whichloose tting keys or props 230` are placed. These props should be fittedloosely-enough that they will slide freely endwise even when theremaining external splines on the jaw members are transmitting torque,and should preferably have a length which will so prop the collar 224and washer 226 apart that the spaces 232 and 234 when added togetherwill equal the maximum totaltravel of the shift collar 224.

An extension-236 of the carrier front member 158 is end splined theretoat 248, the forwardv end of the `extension having outwardly extendingjaws 242 which normally align with and t between corresponding internaljaws 244 in jaw member A bearing bushing 24S is press fitted into theextension and rotatable ony hub B52. End-jaws 248 on the outer end ofthe ring gear front bearing member |18 are adapted to'mesh withcorresponding internal jaws 250 on the inside of jaw member 2 i 2 whenthe clutch shifts to direct drive.

A shift ring 252 fits freely in the inside of jaw member 2 l0 and hasthree legs 254 freely slidable in notches formed in the inside'of thejaw member. The legs 254 are just long enough to touch the shift collar224 when the ring 252 is pressed against the internal jaws 244 of themember 2 I8.

A second shift ring 256 fits freely in the inside of jaw member2l2 andhas three legs 258 freely slidablev in notches formed in the inside ofthe jaw member. The legs 258. are just long enoughV to touch the springwasher 226 when the ring'256 is pressed against the internal jaws of themember 2I2. A shift spring 260 'holds therings 252 and 258 spread apartas shown, altho either ring may be moved toward the other byv pressureagainst the outer ends of the legs 254 or '258 without moving the jawmembers 2li] or 2|2.

A ratchet spring 262 is so placed as to constantly urge the jaw members.2li) and 2|2 axially apart, and, in order to limit their relative axialdisplacement, a means shown more or less diagrammatically in Fig. 17 isemployed. This means consists of hooks 264 formed integral with thebodies of the jaw members whichvallow them to move closer together totake up the spaces 268 and 258, but not spread farther apart.

One pair ofv hooks only is shown forillustration but it will be apparentthat several such pairs of hooks may be circumferentially spaced `aroundthe jaw members. 1 y

The jaws 244 and 259 are beveled on their outer faces as at 218 and 212,while the jaws 242 and 248 are beveled on their inner faces as at 214and- 216, (see Fig. 11). The amount of this bevel should preferably beequal to one-fourth the face width of the jaw. 'Ihe view Fig. 17 istaken looking in the direction of the arrow 218, Fig. 5.

'I'he overdrive clutch is carried on the transmission input shaft 68which has external splines 288. Jaw members 282 and 284 have internalsplines which are axially slidable on the splines 288. Jaw members 232and 284 are exactly alike except that one is turned end for end withrespect to the other. 'I'hey are given different numerals only tofacilitate subsequent description of their operation. On the outside ofthe transmission output member |45, four weights 286 are equally spacedand hinged between ears 288 by hinge pins 298, the work arms 292 of theweights extending thru slots 294 in the output member. The outwardmovement of the weights is accurately limited by contact of the workarms 292 with the front edges of the slots. "Ihe inner ends of the workarms 292 extend into an annular groove in a shift collar 296 which isfreely rotatable on a small forwardly extending hub of the output member|45. A shift washer 298 has internal splines which slidably fit theexternal splines 288 of the shaft.

A weight return spring washer 388 has internal splines also slidable inthe external splines 288, and a weight return spring 382 underconsiderable stress urges the washer 388 rearwardly.

In the instant embodiment there are nine equally spaced external splines288, and on the shift washer 298 and weight return spring washer 388there are nine internal splines slidable on the external splines. In thejaw members 282 andl 284, however, every third internal spline is cutaway (see Fig. 1) which leaves three spaces between external splines 288into which loose fitting keys or props 384 are placed. These propsshould be tted loosely enough that they will slide freely endwise evenwhen the remaining internal splines in the jaw members are transmittingtorque, and should preferably have a length which will so prop thewashers 298 and 3188 apart that the spaces 386 and 388 when addedtogether will equal the maximum total travel of the shaft collar 296.

The rearwardly extending end of the ring gear rear bearing member |88has inwardly extending jaws 3|8 which normally align with and fitbetween corresponding external jaws 3|2 on the jaw member 282. End jaws3|4 on the outer end of the carrier rear bearing member |62 are adaptedto mesh with corresponding external jaws 3|6 on the outside of jawmember 284 when the clutch shifts to overdrive.

A shift ring 3|8 fits freely over the outside of jaw member 282 and hasthree legs 328 freely slidable in notches formed in the outside of theJ'aw member.

The legs 328 are just long enough to touch the shift washer 298 when thering 3|8 is pressed against the external jaws 3|2 of the jaw member 282.A second shift ring 322 fits freely over the outside of jaw member 284and has three legs 324 freely slidable in notches formed in the outsideof the jaw member. The legs 324 are just long enough to touch the springwasher 388 when the ring 322 is pressed against the external jaws 3|6 ofthe member 284. A shift spring 326 holds the rings 3|8 and 322 spacedapart as shown, altho either ring may be moved toward the other bypressure against the outer ends of the legs 328 or 324 without movingthe jaw members 282 or 284.

A ratchet spring 328 is so placed as to constantly urge the jaw members282 and 284 axially apart, and in order to limit their relative axialdisplacement, a means shown more or less diagrammatically in Fig. 8 isemployed. This means consists of the hooks 338 formed integral with thebodies of the jaw members which allow them to move closer together totake up the spaces 332 and 334 but not spread farther apart. One pair ofhooks only is shown, but it will be apparent that several such pairs ofhooks may be circumferentially spaced around the jaw member.

The jaws 3|2 and 3|6 are beveled on their outer faces as at 336 and338,'while the jaws 3|8 and 3| 4 are beveled on their inner faces as at348 and 342 (see Fig. 8). The view Fig. 8 is taken looking in thedirection of the arrow 344, Fig. 7.

Surrounding the transmission output member |45 is an insulating ring 346around which is a metal ring 348. The metal ring 348 has a laterallyprojecting strip 349 which carries a metal contact stud 358 and anupwardly extending strip 352 thru which along screw 354 extends to carryelectric current to a contact lug 356. The lug 356 is imbedded in theinsulating block 35`8 which is held to the head |94 by a screw 368. Thecontact stud 358 is supported in an insulating clamp 362 held to theoutput member |45 by the screw 364. Other insulating clamps 365l arecircumferentially spaced apart to provide additional support for therings 346 and 348. An insulating tube 366 and washer 368 electricallyseparate the screw 354 from contact with parts |94 and |45.

One only of the weights 2|4 carries a plunger 318 backed up by a spring312 which normally keeps the plunger in contact with the block 358 asshown. One only'of the weights 286 carries a plunger 314 backed up by aspring V316 which normally keeps the plunger in contact with theinsulating clamp 362. lWhen the weights 2|4 and 286 swing on theirhinges exactly one-half of their total Voutward swinging movement, theplungers 318 and 314 will be aligned'with the lug 356 and stud 358respectively. A collar 311 is fitted to the outs ide of the ears 288andis slidable Iaxially forward fromthe position shown over theoverdrive weights 286to hold them from moving outward undercertain-circumstances, the purpose .of which will hereinafter be morefully described. Collar 311 is externally grooved to receive theshifting fork 319. i l v A flanged brush holder 318 and itscover 388 ofinsulation are held to the housing-38 by screws 382, and carry a brush384 which is held in electrical contact with the ring 348 by a spring386 which abuts against the head of a binding screw 388 having bindingnuts 398 and washers 392. A flexible lead 394 electrically connects fthebinding screw to the brush. A resistance 319 and wire 38| may conductcurrent from the post 388, to a battery 383, thru a small lamp 385,orother signalling means to indicate when either the plunger 318 or 314has contacted the lug 356 or stud 358 to complete a circuit to theground.

A small bi-,metallic thermostat 381 may be connected in parallel withthe resistance 319 by wires 39| and V393 whereby expansion from heatwill cause the contact 395 to engage the adjustable grounded screw 391.

The long hub |98 of the transmission output member |45 extendsrearwardly into the reversing gear compartment. The reversing sun gear396 has internal splines 398 which t external splines on the hub.

The tail shaft 488 is rotatably supported near the rear end by the ballbearing 402 held against axial movement in the bearing head by the snaprings 404 and at the front end by the bearing bushing 466 which is pressfitted into the rear end of the hub. The larger diameter of the tailshaft 480 'abuts the rear end of the sun gear 396 and thereby preventsthe sun gear from moving axially.

The rear end of the tail shaft is threaded for the nut 408 which holdsthe speedometer driving gear 4l0, the ball bearing 482 and the universaljoint member 412, which has internal'splines 4I4 fitted over theexternal splines on the tail shaft.

A cup 4i6 held in the bearing head 35 contains packing 4l8 which ts theuniversal joint member closely. A speedometer driven shaft 428 withintegral gear 4122 is rotatable in the bearing member 424. The ringgear426 is shown integral with the tail shaft 488 but may be separatelymade and permanently secured thereto.

r"he reversing planet pinion carrier front bearing member 428 isprovided interiorly with the bearing bushing 436 within which the hub ofthe sun 'gear 398 may rotate. Integral hollow hubs 432 extend towardeach other to rotatably support the planet pinions 434 in constant meshwith both the sun gear and the ring gear. The pinions 434 are providedwith bearing bushings 436 which are rotatable on the hubs 432. Thecarrier rear bearing member 438 is held tothe front member 42S by thebolts 448 and nuts k442. A bearing bushing 444 in member 438 isrotatable on the tail shaft. Y

At the forward end, the carrier member 428 has external teeth 446adapted to t slidably into the internal teeth of the plate 448 which isheldto the partition 34 by rivets 450.` Member 428 also has internalteeth 452 adapted to t slidably over the teeth of the sun gear 398. Nearthe forward end, the member 428 is groovedto reive the shifting fork454.

The mechanism for operating the shifting iorks 379 and 454 is carried onthe cover member 456 held to the underside of the transmission housingby screws 458. Parallel guideways are formed in the cover member for theshifting slides 466 and 462 which carry the shifting forks 313 and 454respectively.

Pockets 464 and 466 are formed in the cover 456 directly under theslides 468 and 462 respectively. In the section Fig. 1, the plane of thesection is so deflected as to pass thru the front slide where thesection 9 9 is taken and thru the rear slide where the section l3-l3 istaken.

Pocket 466 contains a shaft arm 468, the upper end of which extends intoa notch in the edge of the slide 462, while the lower end has aninternally splined hub which fits over the external splines 4'ill of arock shaft 472.

Pocket 464 contains a shift arm 414, the upper end of which extends intoa notch in the edge of the slide 468, while the lower end has aninternally splined hub which fits over the external splines 416 of arock shaft 418.

A forward, neutral and reverse lever 480 is tightly secured to the outerend of the rock shaft.

4'52 while an overdrive lockout lever 482 is tightly secured to theouter end of the rock shaft 418. Both rock shafts are provided withconventional packing means comprising an integral collar 484, two bevelfaced collars 486,

packing 488, a spring 499 and a threaded ven cap 492 all as in commonpractice. y

A detent means for each slide (see Fig. 11) comprises a ball 494 andspring 496 in a cup 498. A series of depressions 540 in the bo-ttoms ofthe slides locate the slides in their several positions.

The entire transmission is lubricated by tapping the main oiling systemof the engine, whereby no separate oil pump or reservoir is required.The oil is forced out thru the rear end of the crank shaft 31 into thehollow interior of the shafts 68 and 469, from which it is distributedto various points requiring lubrication lthru radial holes (not shown).f

Centrifugal force is depended upon to carry the oil from the interior ofthe shafts to the bearings, gears, etc., and the oil is preferably notmaintained under pressure. To insure this condition, the holes 582 and584 are provided and so located that one is always open.

The large opening in the centerof shaft 68 is restricted at the frontend by a press fitted bushing 506 whereby some oil is always trappedand. retained in the shaft and is therefore available for starting untila` new supply comes thru from the engine pump.

Since this oil is not delivered under pressure, a packing washer 598effectively prevents leakage of oil between shaft 68 and bushing |43.Holes 5H) transfer any slight amount of leakage oil to the outside ofthe clutch housing. A tapped oilreturn hole 512 in the bottom of thehousing 38 may transfer the accumulated oil. to a lter (not shown) andthence back to the cil reservoir of the engine.

Proportion While this transmission may be proportioned for use with anengine of any horsepower and with vehicle weights within commonpractice, some suggestion as to proportion of the various parts and theprocedure in determining the same for a given case may preferably begiven.

If the diameter of the clutch housing 26 'at section 2-2 is taken as 13%and all parts made to the same scale, the transmission will .be suit-vable for an engine of 110 H. P. at 3600 R.. P. M. in a vehicle weighingapproximately 3500 lbs.

For the reverse gear set Within housing-33 Where quiet operation andlong wear are not the The gearing of the transmission gear set'in thehousing is 14 pitch 14 degree pressure angle, 14 degree helix angle. Thering gear hand should correspond to the threads ina right hand nut. Thering gear H6 has'teethY on ahpitch diameter of 4.196, the sun gear |46has 27 teeth on a pitch diameter of 1.988 and the planet pinions |12have l5 teeth ona pitch, diameter of 1.104".

The ratio through the transmission gear set only, at low speed andbefore either transmission clutch has operated is therefore R+S 57|27 R57 input revolutions to 1 output revolution, the Yratio after the directdrive transmission clutch has operated will, of course, be 1 inputrevolution to 1 output revolution, and the ratio after the overdrivetransmission clutch has operated will be R 57 R14-S 57 +27 inputrevolutions to 1 output revolution.

To provide a well graduated range of speed ratios a booster gear set isnow selected which will have a ratio of 1.6 to l. This ratio may b-e hadwith a ring gear having 60 teeth, a sun gear having 36 teeth, and planetpinions having 12 teeth the ratio being Since the helix angle and thepitch diameters of the booster gears depend on certain factors having todo with the auxiliary dry plate clutch, the proportion of which is notyet, at this time, determined, these tooth angles and dimensions will bedetermined later.

However, since the ratio of the booster gear set has been selected as1.6 to 1, by now selecting the rear axle ratio, the overallengine-to-wheel ratios may be determined. With the engine power andvehicle weight hereinbefore selected and a transmission having anoverdrive, it is average present practice to use a 4.82 to 1 axle and29" wheels.

The booster gear set in itself operates either as an underdrive or adirect drive, while the transmission gear set may be coupled forunderdrive, direct, or overdrive. Therefore the overall engine-to-Wheelratios will be the booster gear ratio the transmission gear ratio theaxle ratio. With these ratios determined the forward engine-to-wheelratios are as follows:

Ratios booster transmission X axle engine-to-wheel The two reverseratios will be as follows:

It has been found expeditious to first determine the dimensions of thedry plate clutch |00 using as large a plate 94 as may be contained inthe space available, whereby less clutch engaging pressure is requiredand therefore a less steep helix angle is required on the booster ringgear II4 to balance said pressure.

In the present embodiment a 11%" O. D. 711% I. D. double faced plate maybe used. From present data available on carrying capacity of dry plates,such a disc, in order to just carry the maximum engine torque of 186 ft.lbs. Without slippage must be clamped between the two metal plates I2and 83 with a pressure of 498 lbs. The centrifugal weight 'I5 andsprings 92 which effect this clamping will be so proportioned that Whenthe weights fall in speed to 2250 R. P. M. the clutch will be just atthe point of failure to any longer carry 186 ft. lbs.

Having determined the size of the dry plate 94 of clutch |00 and theaxial pressure required on it to enable it to just carry 186 ft. lbs.torque, the helix angle of the booster ring gear I I4, which is tooppose the spring pressure tending to engage the clutch, and the pitchdiameter and loading of the gear may be found.

To obtain the proper overlap of 20% to 25% between shift up and shiftdown of the booster mechanism, the axial thrust of the ring gear I I4should now be tentatively selected at from 30% to 35%, say 32% over theminimum clutch engaging pressure of 498 lbs. above found. The thrustYwould therefore be 1.32 498=658 lbs. This should be the axial thrust ofthe ring gear when it is being driven 3600 R. P. M. with the engine atits maximum torque for that speed which is 160 ft. lbs.

Thus a helix angle for the ring gear II4 is to be selected which willprovide a 32% greater end thrust when the engine is rotating 3600 R. P.M. and delivering its -then maximum torque of 160 ft. lbs. V(see Fig.16) than the clutch weights need apply to the dry plate to just carrythe maximum torque of 186 ft. lbs.

With a maximum torque of 160 ft. lbs. at 3600 1 1.6 X 1.474X4.82:11.40:10W

2 1.6 X 1.000 X482: 7.72:torque second `3 1.0 1.474 X482: 7.l2:speedsecond 4 1.6 X .6786 X482: 5.24:torque high 5 1.0 1.000 X482: 4.82:speedhigh 6 1.0 X .6786 X482: 3.27:overdrive Since it is the object in thebooster gear set to use the thrust of the helical teeth of the boosterring gear II4 when under load as a torque responsive means to oppose andovercome the force of the engaging springs 92 of the auxiliary clutchIUI), neither the helix angle nor the pitch diameter of .the ring gear II4 may be determined without reference to the size and capacity of thedry plate clutch |90 to carry torque without slippage.

The helix angle and pitch diameter of the booster ring gear determineits thrust at the maximum H. P. point of the engine, i. e. 3600 R. P. M.and this thrust must be exactly equalled by an opposite force generatedby the centrifugal Weights 'l5 and stored in springs 92 at 3600 engineR. P. M. which transmitted through the 1.6 to 1 booster gear is 2250weight R; P. M.

X 2.0 X4.82:14.24 X 2.() X4.82:22.80

R. P.,M. the pitch line load L on the ring gear II4 at 3600 R. P. M. is

12!! ring gear pitch radius and in order to provide 658 lbs. end thrust,the pitch line load must be tan. helix angle radius which is 2.917Xthetan. of the helix angle of its teeth.

A=1.875 sec. A, or tan A=.64279 sec. A, but since and then bysubstituting sine A .64279 cos. A cos. A

and by cancelling sine A=.64279 whereby A=40.

The helix angle of `the ring gear is therefore 40.

When starting with a dry plate of a given size and working out the ringgear helix angle as above, the helix angle will not always be an evennumber of degrees, but by taking the nearest even number of degrees tothe angle arrived at, and working backwardly through the equations tothe plate, the plate may be slightly modified to correspond to the evendegrees selected.

With the helix angle of the booster gear set fixed at 40, and the pitchand number of teeth known, the remaining dimensions of the gearing willbe as follows: The ring gear has 60 teeth 16 pitch=g pitch diameter fornonhelical teeth vand 60 o 16Xsec 40 for the helical teeth selected. Thepitch diameter of the ring gear is therefore the sun gear t Y -X1.3054=2.937"

and the planet pinionsV Q' Vl/ 16X 1.3054? .979

andthe springs will be applying this 663 lbs.- to Aovercome the gearthrust of 663 lbs. and thereby engage the clutch. The'weights will berevolving 2250 R. P. M. at this time.

vThe diagram Fig. 15 shows the action-Hof the weights 'l5 undercentrifugal force,v the eamount of axial movement of the plate 86, theresulting deflection of the springs 92, the stress of the springs causedby that deflection, and the R. P. M. at which the weights must revolveto produce that stress. p

In Fig. 15, o represents the center of the hinge pin 16; a representsthe center of gravity of a weight 15, as well as the center of a roller19, when the weight is at its home position; h represents the center ofgravity of a weight when it is in the outermost position; b to g areintermediate positions to which the center of gravity may swing in thecourse of its outward swinging movement. Y

The distance ao=ho=.770". The points a to l1. are spaced 10 apart; a. isat the 10l position; h. is at the position; thevertical and hori- Zonta]distances oa, ob, etc. are the sines and cosines respectively of 10, 20,etc., X .770. From the sines, the distances R from the axis of thetransmission to the center of gravity of a weight in all positions a. toh, are found and tabu'- lated. From the cosines, the amounts that thesprings 92 are deected by movement of the center of gravity from anypoint a to h are found and tabulated. The length of the springs, whenthe center of gravity is at a., is 1.375".

Column 1, Fig. 15, shows the axial movement of spring plate 86 for themovements of Weights 'l5 traveling from a to h; column 2 the deflection`ofsprings with weights in their several positions when the clutch |00is engaged; and column 3,

the deflection of the springs with weights at a.

to vh when the Vbooster gear is operative. The reason that the twocolumns have different springdeflections is that, in order to close upthe clutch, the spring plate 86 moves .0915" without chang! ing thespring length'. Therefore when the gear is operative as shown in Fig. 1,the outward swinging of a weight 15 immediately starts 4to shorten thesprings 92, but whentheclutch. is engaged, the weights must swing from ato c to take up the slack between the plates, which is .0915", `Withoutchanging the spring length.

YThese deiiection values may not begotten at this time but must awaitVdetermination of the spring dimensions.

' It has heretofore been the engine is revolving 3600 R. P. M. andtheweights are therefore being driven through the booster gear at 2250 R.P. M., the weights should create a stress of 663 lbs'. in the springs.It must now be determined to what position, a to h, the Weights shallhave swung to be energizingthe springs with a force of 663 lbs. Theposition g', which is the rl0'p0sition, is selected.

Knowing the springs to be 1.3751" long with weights at a, and findingfrom column 1 that .they are shortened .495" in moving to rthel'70"position if the gearris operative, the length in this position Will-be1.375-.495=.880. There are eight springs and theircombined stress is tobe 663 lbs. with weights at g, where the springs are vtabulated incolumn 3 may be determined. vTo

determine the deflections in column 2, the slack .0915" in the clutch issubtracted from each value determined that YvvhenY or S=l39-5 D, where Sis the stress of one spring and D the deflection values of columns 2 and3. For eight springs, therefore, 3:194433.

From the last equation the values for columns 4 and 5 are found fromthose in columns 2. and 3.

Columns 4 and 5 will then show the stress of eight springs in theseveral positions a to h of the weights.

It is now required to so proportion the weights 'l5 that when they arein the g position and are A rotating 2250 R. P. M. they will beproducing an axial force of 663 lbs., so that the springs will be underthis stress in this position and at this Speed.

From Figs. 1 and 15 it will be seen that the farther out a weightswings, the less axial force results per pound of centrifugal force, theradial centrifugal force of the weights in any angular position a to hhaving a resultant axial component of cosine sine of the angle with thetransmission axis the centrifugal force.

The eight weights, therefore, must be of such Weight as will arriveatthe 70 position g where their centers of gravity are 5.0676" from thetransmission axis, when they are rotating 2250 R. P. M., and must bethen generating sufcient centrifugal force to produce` a resultant axialforce of 663 lbs. when the centrifugal force is being translated intoaxial force through a leverage of cos 70 .34202 sin 70 .93969 The wellknown formula for centrifugal force is,

( 1) F=.0000284WRN2 wherein F is the centrifugal `force in lbs., R isthe radius in inches, and N is the R. P. M.

Having found F, the axialV component F' resulting after F has actedthrough the leverage of han .93969 may be expressed in the equation (2)F=.0000103WRN2 which transposed becomes l e cm2 wherein F=663 lbs.,R='5.0676, N=2250 R, P. M. which makes N2=5,062,50O.

Solving, W is found to be 2.500 lbs. for the eight weights. Theirthickness may therefore be varied until the eight weights 75, with theirhubs '18, rollers'IB and washers 8U, will Weigh just 2.500 lbs.

Having found that eight weights totalling 2.500 lbs., rotating 2250 R.P. M., with centers of gravity swung to the 70 position which is 5.0676"from the axis of the transmission, and acting through a leverage of willcreate an axial force of 663 lbs., and that the spring stress is also663 lbs. at this position of the weights, it may now be found at what R.P. M. the weights must be rotating to reach each of the severalpositions a to h and generate forces equal to each of the severaltabulated spring stresses when at the several tabulated distances R fromthe axis, and acting through the several leverages cosine sine of theangles of the several positions.

Repeating Equation 1, F=.0000284WRN2 or F N \/.0000284WR F N- 187.6 W

F=F after it has acted through a leverage of cos A sin A Where A is theangle of any position a to h. So,

cos A (4) sin A ,sin A F-F cos A or F=F tan A.

Rewriting Equation 1 substituting for F the Value of F indicated inEquation 4 We have,

F' ein A 5) N- 187.6 WR

and since W=2.500 lbs.,

wherein A is the angle to which the weight has swung, F=values incolumns 4 and 5, Fig. 15, and R the distances from the transmissionaxis, Fig, 15.

Solving by Equation 6 for the 80 position,

when the clutch isoperative,

N=118.66 5 1023 :3317 R. P. M..

and whenrthe gear isoperative N=118.6c W=sss5 R. P. M.

The values of N are similarly found for all positions a. to h andentered in Fig. 15 columns 6 and 7, from which the force curves t and sof the chart Fig. 16 are plotted. 'I'he curves u and y 'show the forceof centrifugal weights of conventional design. The curve w is themaximum torque curve of the engine selected, and the curverfv a torquecurve less than maximum selected for illustration.

At the topvof the chart Fig. 16 the R. P. M. of the weights 15 from 0 to4200 is given', and the corresponding M. P. H, with the transmissiongear set in underdrive, direct drive, and overdrive are tabulated. Thetwo columns at the left show theengine torque values and thecorresponding axial force which these values generate in the helicalring gear ||4 to hold'the clutch disengaged.

The light spring 5| which acts in favor of the helical gear thrust istaken at 5 lbs. and is added to each value of gear thrust in column 2.This 5 lb. spring should be made of a round wire of S3-2 diameter coiled21/2" pitch diameter, have 2 coils and a free length of 1.781". The twocolumns'at the right show the forcevwhich the weights may store in thesprings 92 and the corresponding foot lbs. carrying capacity of theclutch |00. 'Ihe chart will hereinafter be referred to in describing theoperation of the transmission.

For the main engine clutch 10, the same size dry plate is used as forthe auxiliary clutch |00. If the weights 50 are made to the scaleproposed, and the springs 58 made of .080 round wire coiled pitch,diameter, have 8 coils, and a free length of 1.840, the clutch will havedeveloped a carrying capacity of about 18 ft. lbs. at 450 engine R. P.M. which gradually rises to 180 ft. lbs. lat about 750 engine R. P, M.This makes for a gentle engagement.

The several parts of the direct drive clutch may be made to the scaleselected, but in order to have the clutch shift up from underdrive todirect at 15 M. P. H. when the driving force is released, the springs228, 260, 262, and 312 should be made to exact specifications.

The spring 228 should be made of .156 round wire, coiled a-s" pitchdiameter, have 21/2 coils, and a free length of 4.82 inches. The spring260 should be made of .093" round wire, coiled 2% pitch diameter, have31A? coils, and a free length of 5.45". The spring 262 should be made of.093 round wire, coiled 21/2" pitch diameter, have 7 coils, and a freelength of 3.1".

The spring 312 should be made of .020" round springs 302, 326, 328, and316 should be made to exact specifications.

The spring 302 should be made of .105" round wire, coiledlg pitchdiameter, have 5 coils and a free length of 6.4". The spring 326 shouldbe made of .080 round wire, coiled 1% pitch diameter, have 6 coils and afree length of 4.17". The spring 323 should be made of .080 round wire,coiled 2%" pitch diameter, have 12 `coils and a free length of 5.38".The spring316 is made exactly like spring 312. With the above springs,the transmission shift from direct to overdrive will occur at 30 M. P.H. and the shift back down at 23 M. P. H.

Operation is shown in the drawings, where the centrifugal weights ofthemain clutch 10, the auxiliary clutch |00, and the transmission clutches204 and .206 are all in their "clear in positions and the 're'- versinggear set is in neutral. In this condition the engine may be run andwarmed if desired.

As the engine. speed rises, the clutch 10 first engages and operates thebooster gear, whichin -turn'rotates the shaft 68 which in turn operatesthe clutch |00.` The weights of the transmission clutches also operatein and out at certain points in the rise and fallof Ythe speeds. Thislimbers up not only the .engine .but the entire transmission mechanism.'Nopower is transmitted because the reversing gear is in neutral. i

YTo set the reversing gear set formoving the vehicle backwardly, thelever 480 is drawn forwardlywhich draws Vthe front slide 462 to whichthe'fork 454 is secured forwardly, which moves .the carriermember 428forwardly untilthe teeth v446 meshwith the-teeth of plate 448. When thecarrier isthus heldnon-rotative, forward rotationof the sun gear 306rwill c-ause rearward rotation of ring gear 426 and the Vvehicle willmove vbackwardly.v For all forward driving, the lever 480 is pushedrearwardly whichv pushes .the fork 454 rearwardly `and'slidesthe carrierteeth` 452 over the front ends of the `teeth of the sun gear 396. Theteeth of the planet .pinions 43.4, being still meshed one- -third theirlength into the teeth Iof both the sun gear 396 and ringgear 426, alockedconditionfis provided wherein the tail shaft 400 must rotate inunison with the ltransmission output member |45. 5

l .If the engine is" now speeded up past vr400 R. P..M. the main clutch-10 engages, drives vthe booster vring gear`||4 which starts revolvingthe booster sun gear. |08 backwardly, which is immediately arrested bythe roller brake |01, whereupon the carrier 99 rotates forwardly atreduced speed. Y Y

The carrier is secured to the transmission input shaft 68, thereforeboth rotate at the same speed.V The input shaft 68 being normallyconnected by the clutch 206 to the ring gear element 200 of 4thetransmission gear set While the carrier element 202 is connected by .theclutch.204 to the output member |45, both gear sets will be in seriesand both operating at reduced output member speed. This provides lowgear or rst speed andthe engine to wheel ratio will be 11.4 to 1.

Now as soonas the vehicle starts moving, the auxiliary clutchV |00starts rotating, whereupon the weights 15 start moving out andcompressing the engaging springs 02. There is an unvarying positionYofthe weights and an unvarying length and stress of the springs foranygiven vehicle speed. Whether this stress will engage the auxiliaryclutch |00 or not depends on the forward axial thrust of the -boosterring gear ||4. If this thrust is zero, as for instance when the vehicleis Yallowedv to start itself on a steep down grade, the

`clutch |00 will engage almost immediately following the beginning ofvehicle movement, the light spring-5| being overcome at a' speedof 'lessvthan 1 M. P.'H. This isA important', for it insures en,- gine brakingunder Vany and all circumstances even `when the engine is dead or idlingand the main clutch 10 lis consequently disengaged.

It also permitsthe engine to be started when the batteryis dead bypushing the vehicle.' VOf course, after the engine is rotated thru theauxiliary clutch |00 by vehicle movement to a speedof 400 R. P.M., o rmore, the marin clutch-10 will automatic-ally engage.

If, on the other hand, a start is being made against vehicle resistanceas is substantially valways the case, then the booster ring gear willthrust forward in proportion to that vehicle resistance to opposeauxiliary clutch engagement, and its engagement will be proportionatelydelayed.

lThe booster clutch dry plate, the springs for engaging it, the weightsfor engaging the springs, and the helix angle of the booster ring gearteeth, are all so proportioned, however, that engagement of the clutchand consequent elimination of the booster gear as a factor in reducingspeed and multiplying torque cannot be delayed to the point where theoverall efficiency is too low. The -arrangement is therefore 'such thatno matter what ratio is in effect in the transmission gear-set, ifmaximum torque is being created as represented by the curve w, Fig. 16and kept at maximum until the engine speed reaches 1:3600 R. P. M.: 160foot pounds torque:663 pounds clutch engaging force, the centrifugalweight force will have brought the engaging stress of the clutch springsto 5:2250 R. P. M.:663 pounds clutch engaging force, wherefore the twoforces are balanced. If

the engine speed increases further the engaging force rises and thedisengaging force drops whereby clutch engagement becomes involuntary atthe engines maximum H. P. point, and by engagement it brings the enginefrom 1:3600 R. P. M.: 160 foot pounds torque to k=2250 R. P. M.:18\'J`foot pounds torque.

When the clutch engages at lc:2250 R. P. M. all load is removed from thering gear and there is zero clutch disengaging force. clutch engages,the clutch springs become .0915" longer and therefore weaker, the clutchengaging force dropping from 7:663 pounds to 1:585 pounds which,according to the chart, enables the clutch to carry 220 foot poundsload.

If henceforth the vehicle resistance is such that 'the engine speed canrise above 16:2250, the clutch capacity will rise beyond l on the curvet, but if the vehicle resistance now happens'to be such that the enginespeed is caused to fall from 7c toward n, then the clutch capacity willfall from l toward m.

Assuming the latter condition to be in effect, when the engine speedfalls to n=1750 R. P. M.:

186 foot pounds, the' carrying capacity of the clutch will have fallento m=186 foot pounds, and since the torquelbeing applied and thecapacity of the clutch are now equal, any further reduction in speedwill allow the clutch to disengage and the booster gearing to againbecome effective. The booster ring gear has suiicient axial movementwhen it assumes the load to push the clutch wide open, whereby there isno clutch drag while the gearing is in elect.`

Thus the maximum engine torque kept the clutch |00 fromengaging up to2250 R. P. M. of the `input shaft 68, but when it once engaged, theclutch carried the maximum torque until a reduction in speed of theinput shaft 68 to 1750 R. P. M. occurred. This 23% overlap is desirableto prevent too frequent shifting from clutch togear drive with itsresulting iwear.

The foregoing described Voperation of the booster gear and clutch wasbased on the maximum torque curve w but it will be obvious that whenpower is generated which may be defined by a lesser torque curve, theclutch engaging force, rising along the curve s will meetthe lessertorque curve at a point short of i, the overlap be- When the ing` aboutequal in percentage to the overlap of curve w.

Obviously, by purposely first depressing, then releasing theaccelerator, a torque curve may be created which will drop sharply atany desired speed to a clutch disengaging force line which will coincidewith the line to which the clutch engaging force curve s has risen, andthe clutch will engage. In this way the booster clutch may be engagedand the booster gearing eliminated at will. Further, a sharp depressingof the accelerator when the booster clutch is engaged will raise thetorque curve to a line which passes thru the point `on the curve t whichrepresents the then existing clutch carrying capacity, whereby thebooster clutch may be disengaged at will.

, It will be seen that the booster gear set is controlled by balancingspeed against torque. The transmission gear set is, however, adapted tochange ratios in response to speed only. At 15 M. P. H., the centrifugalweights 2I4 have centrifugal forcesufficient to overcome the weightreturn spring 228 plus the shift spring 260 and consequently the weightsmove out until the shift collar 224 closesthe gap 232 and is stopped bycontact with the clutch member 210.

In closing the gap 2,32, the shift collar 224 has moved the prop 230 andthe ring 252 rearwardly, thereby compressing the springs 228 and 260 anamount equal tothe gap. In closing the gap the weights have swungoutwardly one-half of their total travel which aligned the plunger 310with the lug 356 thereby completing an electric circuit thru the signalshowing that a shift ought now to be `made from underdrive to direct.

No shift will be made, however, as long as torque is being transmittedthru the clutch member 2l6, for altho the weight 2l4 is attempting tomove thru the other half of its outward travel, it cannot do so againstthe frictional resistance in the splines 208 and jaws 244 under load,and while the spring 260 has now been compressed by the ring 252 to urgethe clutch member 212 rearward, it cannot move because it is held to theclutch member 2 I 0 by the hooks 264 (see Fig. 17)

As soon as the signal, that a shift is possible, appears, the operatormay momentarily release the applied torque by letting up on theaccelerator pedal, whereupon the weights 214 will move outwardly theother half of their travel, thus pushing the clutch member 2I0rearwardly far enough that its jaws 244 are slightly more than half outof mesh with the jaws 242 whereby, because of the beveled faces 210 and214, the engine speed may quickly fall, the jaws 242 revolving slowerthan the jaws 244 by ratcheting over them.

When the jaws 244 moved far enough axially to be slightly over-,half outof mesh with the jaws 242, the hooks 264 permitted the jaws 256 to moveaxially far enough to be slightly less than half in *mesh* with the jaws248, altho the spring 260 has now beenstressed Ysufficiently by movementof the ring 252 to bias the jaws 250 for full depth entry into the jaws248.

When thev hooks 264 first allowed the spring biased jaws 25E) to try toenter the ring gear jaws 243, the ring gear -jaws 248 were revolving toofast for engagement, but the bevels 212 and 276 are such that the jaws248 merely ratchet over the jaws 250.

'During the rst second of this shift up period, when both sets of jawsare thus ratcheting, and vthe clutch members 2H). and 2I2` are beingrevolvedY by vehicle momentum, the engine speed may be raised, ifdesired, to provide a one way .three seconds, the engine speed willhavedropped to the point where the jaws 248y have slowed down to the speedof the jaws 255, whereupon the spring 268 will slide the jaws 250 intofull depth mesh with the jaws 248 and in doing so cause the hooks 264 towithdraw the jaws 244 completely out of mesh with the jaws 242.

This establishes a positive two way drive between the ring gear |16 andoutput member |45 and leaves no drive between the carrier and the outputmember as previously existed.

When the weights 2| 4 movedall the way out, the circuit was broken atthe point where the plunger 310 moved off of the lug 356.` At this timethe engine speed had not yet been lowered to a point where the secondhalf of the shift was completed. The signal, however, would now havedisappeared prematurely, .that is as soon as ratcheting began and oneway drives were established, had it not .been for the thermostat 381,which expands to complete a ground connection for the signal thru screwl391V several seconds after one is completed by the weights and holdsthe connection several'seconds after the weights break it. The screw, 391 should preferably be adjusted to hold the ground connection for aboutthree seconds after the connection thru lug 356 is severed. V

Inasmuch as the input shaft 68 is normally connected to the ring gear|16 thru the overdrive clutch member 282 and the output member 45 hasnow been connected to the ring gear thru the member 2|2 of the directdrive clutch, the transmission gear-set is in direct drive, lthe ringgear rotating at input member and output member speed and the carrieridling at a lesser speed.

From the foregoing, it will be seen that the shift up of thetransmission gear-set from underdrive to direct involves starting with atwo way connection thru underdrive and none thru direct, then effectinga one way connection thru underdrive and a one way connection thrudirect, then eecting a two way connection thru direct and none thruunderdrive.

It will also be seen that the hooks 234 prevent the clutch 284 makingYal two way driving connection of the output member |45 with the fasterring gear and the slower carrier at the same time. Obviously, as betweenthe ring gear and the oarrier, either may have positive two wayconnection with the output member as long as the other has only aratchet connection. vThis condition prevents freewheeling inasmuch asthere never is a point in the transition period whenthe engine will notdrive the vehicle and the vehicle will not drive the engine, Y

From the fact that the shift up of the direct drive clutch was possiblewhen the centrifugal force of the weights 2| 4 exceeded the stress ofthe weight return spring 228 plus the shift spring 268, it will beunderstood that a shift back down becomes possible when the spring 228exceeds the centrifugal force of the weights 2 |4 plus the shift spring268. M. P. H., the vehicle speed must be reduced to about 11.8 M. P. H.,to cause the weight to lose an amount of centrifugal force equal todouble the stress'of the shift spring 260 before a shift down willoccur. The shiftdown will then be After a shift up, therefore, at

accomplished by repeating the shift up process except in the oppositedirection.

At 30 M. P. H., the centrifugal weights 286 have centrifugal forcesufficient to overcome the weight return spring 382 plus the shiftspring 326 and consequently the weights move out until the shift collar296 and washer v298 close the gap 386 and are stopped by contact withthe clutch member 282.

In closing the gap 386, thewasher 288 has moved the prop 384 and thering 3|8 forwardly, thereby compressing the springs 382 and 326 anamount equal to the gap. In closing the gap the weights have swungoutwardly one-half of their total travel which aligned the plunger 314with the contact stud 35|) thereby again completing an electric circuitthru the signal showing that a shift ought now to be made from direct tooverdrive.

No shift will be made however, as long lastorque is being transmittedthru the clutch member 282, for altho the weight 286 is attempting t0move thru the other half of its outward travel, it cannot do so againstthe frictional resistance of the splines 280 and jaws 3| 2 under load,and while the spring 326 has now been compressed by the ring. 358 tourge the clutch member 284 forward, it cannot move because it is held tothe clutch member 282 by the hooks 330'A (see Fig. 8).

As soon as the signal, that a Vshi-ft is possible, appears the operatormay again momentarily release the applied torque, whereupon the weights286 will move outwardly the remaining half of their travel, thus pushingthe clutch member 282 forwardly far enough that its jaws 3 2 areslightly more than half out of mesh with the jaws 3|0 whereby, becauseof the beveled faces-336 and 348, the engine speed may be let down, thejaws 3|2` revolving slower than the jaws 3| 0 by ratch- `eting overthem.

W'hen the jaws 31|2 -moved far enough axially to be slightly over halfout of mesh with the jaws 3|8, the hooks 338 Ipermitted the jaws SIB tomove axially'far enough tof be slightly less than half in mesh with thejaws 3|4, altho the spring 326 has now been stressed sufficiently bymovement of the ring 3|8 to bias the jaws 3| 6 for full depth entryintothe jaws 3|4. v

When the hooks 338 rst allowed the spring biased jaws 3|6 to try toenter the carrier jaws 3|4, the carrier jaws 3|4 were revolving too slowfor engagement, but the bevels 338 and 342 are such that the jaws 3|3merely ratchet over the jaws 31|4.

During the first second of the shift up period, when both sets of jawsare thus ratcheting, and the clutch members 282 and 284 are beingrevolved by engine momentum, the engine speed may be raised, if desired,to provide a one way drive between the jaws SI2 and 3||l whereby thevehicle is still driven in direct drive. This would be done only in anemergency. If, on the other hand, the operator waits longer, preferablyabout three seconds, the engine speed will have dropped to a point wherethe jaws`3| have slowed down to the speed of the jaws 3| 4, whereuponthe spring 328 will slide the jaws 3|6 intofull depth mesh with the jaws3|4 -and in doing so cause the hooks 330 to withdraw the jaws 3|2completely out of mesh with the jaws 3|8. This establishes a positivetwo way drive between the carrier member |82 land input shaft 68 andleaves no drive between the ring gear and the shaft 68 as previouslyexisted.V

When the weights 286 moved all the way out.

current was broken at the point where the plunger 314 moved off theContact stud 350. The thermostat 381 again holds the signal in effectuntil ample time has elapsed for the shift to be completed.

Inasmuch as the input shaft 68 is now connected to the carrier thruclutch member 284 and the output member |45 is still connected to thering gear, thru clutch member 2 l2, the transmission gear-set is now inoverdrive, the output member rotating faster than the input member.

From the foregoing it will be seen that the shift up from direct tooverdrive involves starting with a two way connection thru direct andnone thru overdrive, then eifecting a one way connection thru direct anda one way connection thru overdrive, then effecting a two way connectionthru overdrive and none thru direct.

It will also be seen that the hooks 330 prevent the clutch 206 making atwo way driving connection of the input shaft 68 with the faster ringgear and slower carrier at the same time.

Obviously, as between the ring gear and the carrier, either may havepositive two way connection with the input shaft as long as the otherhas only a ratchet connection. This condition prevents freewheeling,inasmuch as there never is a point in the transition period when theengine will not drive the vehicle or the vehicle drive the engine.

From the fact that the shift up of the overdrive clutch was possiblewhen the centrifugal force of the weights 286 exceeded the stress of theweight return spring 362 plus the shift spring 326 it will be understoodthat a shift back down becomes possible when the spring 362 exceeds thecentrifugal force of the weights 286 plus the shift spring 326.

After a shift up therefore, at 30 M. P. H., the

vehicle speed must be reduced to about 23 M. RH., to cause the weightsto lose that amount of centrifugal force equal to double the stress ofthe shift spring 326 before a shift down will occur. The shift down willthen be accomplished by repeating the shift up process except in anopposite direction.

The manner in which the booster gear set c0- operates with thetransmission gear-set to meet different driving conditions may best beillustrated by several examples, reference being had to the chart Fig.16.

Whenever the vehicle comes to rest, the booster gear-set and thetransmission gear-set are automatically connected in series and theengine to A wheel ratio is 11.4 to 1, which is low gear.

As a first example, assume that the driver wishes to bring the vehiclefrom rest to its maximum speed in as short a period as possible. He willof course, create the maximum torque curve w. At 15 M. P. H., the enginereaches a=1978 R. P. M. and the transmission input shaft 68, beingdriven thru the booster gear set at 1 to 1.6 ratio will be revolvingb=1236 R. P. M. and the transmission weights 214 will be revolving c=838R. P. M., the speed at which the weights start moving out.

At this point, 15 M. P. H., the signal indicates that the transmissiongear-set should be shifted fromunderdrive to direct by releasing theapplied power. Let us assume the driver ignores the signal and maintainsthe torque at maximum. When the engine reaches a speed i=3660 R. P. M.,the booster gear weights I5 will be at j=2250 R. P. M., the speed withtransmission gear-set in underdrive being now about 27 M. P. H.

At z' the engine torque has dropped to 160 foot pounds and the helicalgear thrust clutch disengaging force is 663 pounds. The booster weightsl5 have meantime increased the stress of springs 92 along the curve suntil at y' the stress is 663 pounds. The slightest further advance inengine speed will therefore lower the disengaging and raise the engagingforce of clutch |00 and it will engage. I

When clutch |00 thus engages at 27 M. P. the engine speed and torquewill be altered from 23600 R. P.Y M :160 foot pounds torque to lc=2250R. P. M.=186 foot pounds torque, the engine speed and weight speed beingnow equal, and, since the drive is now in direct thru the boostergear-set, and in underdrive thru the transmission gearset, theengine-to-wheel ratio is 7.12 to 1, hereinbefore designated as speedsecond.

If the operator still has not released the applied power, the enginewill again rise in speed from lc=27 M. P. H., to i=44 M. P. H., still inspeed second. Further changes in ratio may now be made only by releaseof the applied torque. If this is done the torque will fall from i toi1, and the engine speed will begin to drop. If the engine speed isallowed to drop from 3600 R. P, M. only as far as e1=2443 R. P. M. andthe power is then reapplied and the torque brought to e, thetransmission gear set will have shifted to direct, and, the boostergear-set being now in direct, the engine-to-wheel ratio will be 4.82 tol hereinbefore called speed high.

By now raising the engine speed from 6:2443 R. P. M. to i=3600 in speedhigh ratio, the vehicle speed is raised from 44 M. P. H., to 65 M. P.H., then by againreducing the engine speed to 61:2443 R. P. M. andreapplying the power so as to bring the torque to e, the transmissiongear-set will have shifted to overdrive and by bringing the engine speedto 3600, the vehicle speed Will be 96 M. P. H.

If, however, the engine speed had been allowed to drop from i to f1=1676R. P. M., before power was reapplied to bring the torque to f, thetransmission gear-set would have shifted from underdrive directly tooverdrive where the engine-towheel ratio is 3.27 to 1. If after suchchange the engine speed were raised to 3600 R, P. M., the vehicle speedwould be 96 M. P. H.

The foregoing first example was given to show that, altho a driver mayignore the shift up signal, the result is not serious, the effect beingthat he may be speeding his engine faster than necessary to attain hisdesired result.

As a second example, to show how, for a similar desired result, theshifts are preferably made to keep the engine operating at the lowerspeeds where its torque is higher, assume that the second driver alsowishes to bring the vehicle speed from rest to maximum in as short atime as possible. He will also create the maximum torque curve w. Whenthe engine speed reaches a=l978 R. P. M., the transmission input shaft66, driven thru the 1.6 to 1 booster gear ratio will revolve b=1236 R.P. M., and the transmission weights 2|4 now revolved thru thetransmission gear ratio 1.474 to 1, will be turning c=838 R. P. M., andthe signal will indicate that a shift up should be made.

The driver now releases the accelerator, and the torque drops from a toai- Since there is now no axial thrust in the booster ring gear H4, the

booster gear set immediately shifts up by en-y gagement of clutchY andassists slightly in re-V to bring it from zero at c1 to maximum at c,the` booster clutch |00 disengages justas the-rising torque crossesthecurve t, and instead of the engine speed and torque arriving at c, itarrives at b, where the engine speed is 1236,/altho the transmissioninput shaft {S8-*still revolvesv c=838 R. P. M.

The booster set being now in gear and the transmission set in direct,the overall ratio is 7.72 to 1 which has been designated ltorque second.This shift was made at M. P.' H.

With torque second ratio in effect, the engine speed rises from b=l236.to :1:2472 R.'P. M. whereby the vehicle speed is `brought Vfrom 15 to30 M. P. I-I., at Vwhich speed the overdrive weights 286 revolve 1676 R.P. M. and move out` Due to lack of torque reaction of the booster' ringgear, the booster clutch .now engages land assists slightly in reducingthe engine speed. By pausing for about three seconds, the vengine speedwill have dropped from d1=2472 R. P. M to g1=1138, the speed at whichthe input shaft 68 must be revolving `to drive the .overdrive .weightsat 1676 R. P. M. thru the overdrive ratio.

When, therefore, the enginel speed reaches 1138, the transmissionoverdrive clutch 20.6 will engage and the signal will disappearfor thesecond time.

The booster gear set and the transmission gear set are again both indirect for an instant, but when maximum engine torque is now reapplied,to bring it from zero at g1 to maximum at y, the booster clutch lill!again disengages just as the torque crosses the curve t, and instead ofthe engine speed and torque arriving at g, it will arrive at h, wherethe engine speed is 1820 R. P. M., but the transmission input shaft 6Bwill still be at g=1138 R. P. M.

The booster gear set being now in gear and the transmission gearing inoverdrive, the overall ratio is 5.24 to 1 which has been denominatedtorque high. This shift occurs at 30 M. P. H.

With torque high ratio in effect, the engine speed rises from h=1820 toi=3600, whereby the vehicle speed is brought from 3Q M. P. I-l., to 59M. P. H., and the booster weights 'l5 will be revolving i=2250 R. P. M.

At the slightest increase in engine speed,rthe booster clutch engagingforce rises above :lL-.663 pounds and the booster ring gear disengagingforce falls below 1":663 pounds and the booster clutch engages,whereupon the engine speed will or 1.474--times as fast as theengin'e,whiclris If the engine speed is'again raised from 2250 to 3600 with theoverdrive ratio in effect, the vehicle speed will again be raised to 96M. P. H. i.

The two 'examples given pertained to the use of the transmission formaximum acceleration,

and it should be noted that in the second example the average enginespeed was lower. It will be seen, however, that even in the secondexample, after each shift up in the transmission gear set,

the booster gear was drawn into series with the' transmission gearbecauseV of the reapplication of maximum torque, whereby the dry plateclutch was twice disengaged and reengaged .in bringing thespeed from 0to 96 M. P; H., which equals the practice in conventional hand shifttransmis` sions with foot clutches.

The cases, however, where it is desirable that the vehicle be broughtfrom rest to maximum speed in the fewest possible seconds are rare, so,in order to show'the better practice in bringing the vehicle from resttomaximum speed, a third example is given as follows: f

The driver creates a torque curve 1J which at its highest point is butfoot pounds, or about half the maximum capacity of the engine.

After a slight spurt in` low gear, he allows the curve to falljoifito q,where the v.torque is about 55 foot Vpounds and the engine `speed 944 R.P; M.

The booster weight 15 now revolves p=590 andthe vehicle speed is 7 M. P.H.

- Since 131:230 pounds=the clutch engaging force of springs 92 and q=230pounds=the clutch disengaging force' of ,booster ring gear H4, thebooster clutch, Ill!) willengage and cause the engine 'speed to lbeloweredfrcm 944 to 590 whichV is the weight speed. `If the torque hasbeen further dropped, say to about 26 foot pounds, engagement of clutchIlllland elimination of the booster gearwill 'occur'at r with 26 footpounds torque still applied, the transmission input shaft 68 revolving590 R. P. M; and the vehicle speed at 7 M.P. H. At'theinstantofengagement of clutch l0!J,'the force fof the springs92-droppedV from 13:23!) pounds to =172 pounds, at which pressure thjecarrying capacity of the clutch Illll is still about -65 foot pounds,which is more than one-third'of themaximum torque'of the engine.

, Since the booster gear set isin direct drive and the transmission gearset in underdrive, thewhereupon the transmission clutch'erigagesandthemsignval disappears. Since the booster' gear set is in direct driveand the vtransmission gear set also Vinv direct 'drive'. the overallratio is 4.82 to l which has been ycalled speed high. The torque is nowreappliedand brought to c2, about 52 foot pounds thenrraised along theltorque curve v to f2' 1676 where the'torque is'l07 foot pounds, thevehicle speed in direct is 30 M. P. H., whereupon thesignal for a shiftto overdrive appears.

By` again releasingthen accelerator so vthe torque win dropfromf2"=1o7foot pounds to' fi`=0, lthen allowing the engine speed to drop to.6786 1676=g2=1138 R. P. M., the transmission overdrive weights 286 willmove out and the signal will disappear and the mechanism will be inoverdrive with overall ratio of 3.27 to 1 and with engine speed 92:11.38R. P. M. and vehicle speed 30 M. P. H.

After the shift up the engine speed may be raised from. 1138 to 3600,bringing the vehicle speed from 30 to 96 M. P. H. in overdrive, withoutonce disengaging the dry plate clutch IDU. Thus in this third examplethere was one engagement and no subsequent disengagement of the clutchin bringing the vehicle from rest to 96 M. P. H.

By operating his accelerator to create the torque curve v, it may besaid that the driver first applied a spurt of half the engines maximunipower in low gear to start the vehicle, then let the power trail offuntil at 7 M. P. H. only about one-sixth maximum engine power was beingapplied, after which he so gradually increased the engine power that atno time did it rise above the curve t, whereby he raised the vehicle'speed in second gear from 7 to 15 M. P. H., then by letting up on theaccelerator pedal he shifted to high, then raised the vehicle speed from15 to 30 M. P. H. in high, being careful not to apply torque above thecurve t, then again let up on the accelerator to shift from high tooverdrive, then 'continue in overdrive to 96 M. P. H.

The reason for the novel spring and weight arrangement of the clutch |00will now be more clear. When centrifugal weights are arranged as inVgeneral practice, the force which they will produce, to engage aclutch, increases with the square of the R. PQM. Such force curves areshown at u and y, Fig. 16.

If there should be created a clutch engaging force :663 pounds at 2250R. P. M. to just balance a clutch disengaging force z' at 3600 R. P. M.,as hereinbefore stated to be a requirement, with weights of conventionaldesign, the curve u would represent the force available at various R. P.M. to keep the clutch from engaging and the curve y would represent theforce available at various R. P. M. to keep the clutch engaged once ithad engaged. It is evident that if curves u and y were substituted for sand t respectively, the performance illustrated by the torque curve vwould be impossible, for most of the curve v lies above instead of belowthe curves u and y.

Thus, when at '7 M. P. H. the driver had dropped his torque to r andthereby eliminated the booster gear in favor of the booster clutch, hecould have raised the torque to, but not above, =42 foot pounds withoutputting booster gear back into play, but had his Weights beenconventional as expressed by curves u and y, he could have raised historque to, but not above, 8 foot pounds without bringing his boostergearing back into play.

The need for the special booster weight and spring arrangement is nowapparent, for it is quite essential that a reasonable amount of torquemay be applied at the lower vehicle speeds without constantly pulling agear set back into play when the operator has no desire for rapidacceleration, nor any hill climbing to do, and therefore no need of thetorque multiplying gears.

The curve o represents substantially the average practice inaccelerating vehicle speed, i. e.,

by raising the torque gradually as the vehicle responds. If, however, atany time during this gradual acceleration, the need for more rapidacceleration suddenly arises, a spurt of power which will throw anupward loop in curve o high enough to rise up and cross the curve t willbring in the booster gear and reduce the ratio from overdrive to high,from high to seoond, or, from second to low as the case may be. Thisupward loop in curve 'u must,l however, cross the curve t aheadof thepoint m, that is, if driving in second gear faster than 22 M. P. H., thebooster gear may be no longer brought into play to provide low gear, andif driving in high gear faster than 32 M. P. H., the booster gear may nolonger be brought into play to provide second gear, and if driving inoverdrive faster than 47- M. P. H., the booster gear may no longer bebrought in to provide high.

The reason for the above arrangement is, that when driving in overdrive,for instance, at 47 M. P. H., the input shaft 68 and the engine will berevolving m=1750. If, now the booster gear could be brought back intoplay the engine speed would have to rise to 1.6 1750=2800 R. P. M. Sincethis is within 22% of the maximum H. P. point of 3600 R. P. M., a shiftup after this speed would not be justified for the reason, that thefalling off of the torque and the time and wear incident to the shiftwould not justify making it.

From the foregoing it will be seen that a high to overdrive shift may beeffected at 30 M. P. H., but overdrive back to high not after 47 M. P.H.; second to high at 15 M. P. H., but high back to second, not after 32M. P. H., while a shift from low to second may be made anytime afterabout 1 M. P. H., but from second back to low not after 22 M. P. H. Thisis as it should be, for there is no advantage in shifting back down to alower ratio, after a vehicle speed has been attained which would requiretoo high an engine speed thru the lowered ratio to effect any gain.

It will of course be understood that in shifting from overdrive back tohigh as above stated, the overdrive clutch in the transmission housing,being responsive to speed only, does not shift back down at 4'7 M. P.H., but retains its overdrive speed increasing gear connection while thebooster gear set draws down into its speed -reducing gear connectionwhich reduces the overall ratio thru both sets to a ratio not far fromthe 1 to 1 ratio had when both gear sets are in direct drive. The sameis true when any other gear ratio is reduced by a torque applicationwhich brings in the booster gear.

In driving thru mountainous country, the overdrive lock out lever 482should be drawn top forward, by any suitable dash control knob wherebythe collar 311 slides over the weights 286 and prevents any shift fromhigh to overdrive. This is more easily done before the weights havemoved out to make the overdrive connection, but may be done thereafterwithout harmful result. Both driving and engine braking are improved bylocking out the overdrive in mountain driv- 1ng. i

I n driving in congested trafhc the overdrive lock out may preferably,tho not necessarily, be used. In zones where a 15 M. P. H. limit isimposed, and where frequent stops at traffic signals are required, anoperator may preferably accelerate from 0 to 15 M. P. H. in low gear,then shift directly to high, as is now done by expert drivers with handshift transmissions. This is

